Series bearing support apparatus for a gas turbine engine

ABSTRACT

A bearing support apparatus for a gas turbine engine, including: a first annular member; a second annular member disposed within the first annular member; first and second axially spaced-apart rolling-element bearings disposed between the first and second annular members so as to enable relative rotation of the first and second members; a first spring element mounting the first bearing to the first member; and a second spring element mounting the second bearing to the first member through the first spring element, such that the first and second bearings are physically supported in series relative to the first member.

BACKGROUND OF THE INVENTION

This invention relates generally to rolling element bearings and moreparticularly to the mounting of such bearings in gas turbine engines.

A gas turbine engine includes one or more shafts which are mounted forrotation in several bearings, usually of the rolling-element type. Thebearings are enclosed in enclosures called “sumps” which are pressurizedand provided with an oil flow for lubrication and cooling. The bearingsin a gas turbine engine are usually a combination of roller and ballbearings. The roller bearings react radial loads of the shafts and theball bearings react radial and thrust loads. Commonly the bearings aremounted in pairs, with a ball and a roller bearing at axially adjacentpositions in a single bearing sump.

The ball bearings are typically the least reliable bearings in anengine. Ball bearings are more susceptible for failure when subjected toa combination of axial and radial load, while roller bearings that areunder-loaded in the radial direction often fail by roller skiddingdamage. Increased radial load on the roller bearings increases cagespeed, which decreases the risk of such damage. Accordingly, suchbearing configurations require isolation of loads, with thrust loadsbeing directed to the ball bearings and radial loads being directed tothe roller bearings as much as possible. Prior art bearing supportdesigns only isolate the majority of radial load.

Some existing engines have addressed the load isolation problem byincorporating “parallel” spring finger housings to support adjacent balland damper roller bearings. This design is not completely effective atisolating the radial load from the ball bearings, with typically 10-30%of the applied radial load being transmitted through the ball bearings.

Other known existing engines address load isolation by including a ballbearing spring finger in “parallel” with a preloaded uncentered damperroller bearing. This design is effective at load isolation but iscomplex to manufacture and operate.

Accordingly, there is a need for a bearing support for a ball-and-rollerbearing combination which isolates radial and thrust loads to theappropriate bearing and which is mechanically simple.

BRIEF SUMMARY OF THE INVENTION

This need is addressed by the present invention, which provides abearing support apparatus in which a ball and roller bearing are mountedin series by spring elements, so as to isolate radial loads from theball bearing.

According to one aspect of the invention, a bearing support apparatusfor a gas turbine engine includes: a first annular member; a secondannular member disposed within the first annular member; first andsecond axially spaced-apart rolling-element bearings disposed betweenthe first and second annular members so as to enable relative rotationof the first and second members; a first spring element mounting thefirst bearing to the first member; and a second spring element mountingthe second bearing to the first member through the first spring element,such that the first and second bearings are physically supported inseries relative to the first member.

According to another aspect of the invention, a bearing apparatus for agas turbine engine includes: a stationary, annular frame; a firstbearing cage comprising annular forward and aft rings interconnected byan annular array of axially-extending first spring fingers, wherein theforward ring is mounted to the frame; a second bearing cage comprisingannular forward and aft rings interconnected by an annular array ofaxially-extending second spring fingers, the aft ring mounted to the aftring of the first bearing cage; a roller bearing carried by the aft ringof the second bearing cage; a ball bearing carried by the forward ringof the second bearing cage; and a shaft mounted in the roller and ballbearings.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention may be best understood by reference to the followingdescription taken in conjunction with the accompanying drawing figuresin which:

FIG. 1 is a half-sectional view of a gas turbine engine incorporating abearing support apparatus constructed according to an aspect of thepresent invention;

FIG. 2 is an enlarged view of a portion of the gas turbine engine ofFIG. 1, showing the bearing support apparatus;

FIG. 3 is a half-sectional view of a portion of gas turbine engine,showing an alternative bearing support apparatus;

FIG. 4 is a half-sectional view of a portion of gas turbine engine,showing another alternative bearing support apparatus disposed betweentwo shafts; and

FIG. 5 is a half-sectional view of a portion of gas turbine engine,showing yet another alternative bearing support apparatus disposedbetween two shafts.

DETAILED DESCRIPTION OF THE INVENTION

Referring to the drawings wherein identical reference numerals denotethe same elements throughout the various views, FIG. 1 depicts a gasturbine engine 10. The engine 10 has a longitudinal axis 11 and includesa fan 12, a low pressure compressor or “booster” 14 and a low pressureturbine (“LPT”) 16 collectively referred to as a “low pressure system”.The LPT 16 drives the fan 12 and booster 14 through an inner shaft 18,also referred to as an “LP shaft”. The engine 10 also includes a highpressure compressor (“HPC”) 20, a combustor 22, and a high pressureturbine (“HPT”) 24, collectively referred to as a “gas generator” or“core”. The HPT 24 drives the HPC 20 through an outer shaft 26, alsoreferred to as an “HP shaft”. Together, the high and low pressuresystems are operable in a known manner to generate a primary or coreflow as well as a fan flow or bypass flow. While the illustrated engine10 is a high-bypass turbofan engine, the principles described herein areequally applicable to turboprop, turbojet, and turboshaft engines, aswell as turbine engines used for other vehicles or in stationaryapplications.

The inner and outer shafts 18 and 26 are mounted for rotation in severalrolling-element bearings. The bearings are located in enclosed portionsof the engine 10 referred to as “sumps”. FIG. 2 illustrates a portion ofone sump of the engine shown in FIG. 1. The inner shaft 18 is surroundedby a stationary structural frame 28 having an annular flange 30 whichextends radially inward.

A first bearing cage 32 is mounted to the flange 30. The first bearingcage 32 can be considered a spring support or a resilient support. Thefirst bearing cage 32 comprises an annular forward ring 34, and anannular aft ring 36, interconnected by an annular array of spring arms38. The spring arms 38 are configured to be essentially rigid againstaxial deflection of the aft ring 36 while permitting controlled radialdeflection of the aft ring 36 relative to the forward ring 34. Thenumber, shape, spacing, and dimensions of the spring arms 38 may bemodified to suit a particular application, in particular to achieve adesired radial stiffness of the first bearing cage 32. As used hereinthe term “stiffness” refers to the force or unit load required per unitdeflection. The aft ring 36 is generally cylindrical and extends axiallyin-line with the spring arms 38. The forward ring 34 extends the radialdirection so as to define a mounting flange. The forward ring 34includes a plurality of mounting holes 40 which receive fasteners 42that secure the forward ring 34 to the flange 30 of the frame 28.

A second bearing cage 44 is mounted to the aft ring 36 of the firstbearing cage 32. The second bearing cage 44 comprises an annular aftring 46 received in the aft ring 36 of the first bearing cage 32, and anannular forward ring 48 positioned forward of the flange 30,interconnected by an annular array of spring arms 50. The second bearingcage 44 is secured to the to the first bearing cage 32 so that it willnot move relative thereto. This may be accomplished, for example, by aninterference fit, by the use of welding, radial pins, a mechanicaljoint, or other similar techniques. The spring arms 50 are configured tobe essentially rigid against axial deflection of the forward ring 48while permitting controlled radial deflection of the forward ring 48relative to the aft ring 46. The number, shape, spacing and dimensionsof the spring arms 50 may be modified to suit a particular application,in particular to achieve a desired radial stiffness of the secondbearing cage 44. In some applications, the radial stiffness of thesecond bearing cage 44 may be selected to be significantly less than thefirst bearing cage 32.

A roller bearing 52 is disposed between the frame 28 and the inner shaft18. The roller bearing 52 includes an annular inner race 54, a row ofgenerally cylindrical rollers 56, an annular outer race 58, and aretainer 60. The inner race 54 of the roller bearing 52 is secured tothe inner shaft 18 so that it will not rotate relative thereto, and theouter race 58 of the roller bearing 52 is secured to the aft ring 46 ofthe second bearing cage 44 so that it will not rotate relative thereto.The inherent configuration of the roller bearing 52 is such that axialloads cannot be transmitted through the roller bearing 52, other thanminor forces transmitted by friction between the rollers 56 and theraces 54 and 58.

A ball bearing 62 is disposed between the frame 28 and the inner shaft18, axially forward of the roller bearing 52. The ball bearing 62includes an annular inner race 64, a row of spherical balls 66, anannular outer race 68, and a retainer 70. The inner race 64 of the ballbearing 52 is secured to the inner shaft 18 so that it will not rotaterelative thereto, and the outer race 68 of the ball bearing 62 issecured to the forward ring 48 of the second bearing cage 44 so that itwill not rotate relative thereto. The inner race 64 defines a concaveinner seat 72, and the outer race 68 defines a concave outer seat 74,both of which overlap the balls 66 in the radial direction. The innerrace 64 abuts bearing surfaces 76 of the inner shaft 18 (or componentsmounted thereto) and the outer race abuts bearing surfaces 78 of thesecond bearing cage 44 (or components mounted thereon). Thisconfiguration permits axial loads (also referred to as thrust loads) tobe transmitted from the inner shaft 18 through the ball bearing 62 andthe first and second bearing cages 32 and 44 to the frame 28.

The bearing support apparatus includes an optional annular damper arm80. The damper arm 80 has a forward portion 82 which is secured to theflange 30, for example using the illustrated bolts 84 or other suitablefasteners. The damper arm 80 extends radially inward and axially aftfrom the flange 30. An aft portion 85 of the damper arm 80 is generallycylindrical and is mounted in axial alignment and close radial proximityto the aft ring 36 of the first bearing cage 32.

Together the outer surface of the aft ring 36 and the inner surface ofthe aft portion 85 of the damper arm 80 define a thin annular squeezefilm space 86 therebetween. A damper fluid such as an oil underpressure, is introduced into squeeze film space 86 through anappropriate conduit or inlet (not shown). In accordance with knownprinciples, any rotor imbalance may cause the inner shaft 18 and theroller bearing 52 to undergo radial motion and subject oil in thesqueeze film space 86 to very high pressure, thereby forcing viscousflow of the oil and a damping action on the roller bearing 52 and innershaft 18. This action limits deflection of the aft ring 36 and theroller bearing 52. In order to seal damper oil in squeeze film space 86,the aft ring 36 includes a pair of spaced apart grooves 88 in its outersurface. A pair of sealing rings (not shown) may be fitted in thegrooves 88 and adapted to engage the surface of the damper arm 80,thereby sealing the forward and aft boundaries of the squeeze film space86.

In operation, the inner shaft 18 is subject to movement in the radialdirection relative to the frame 28, causing radial deflections and loadsin the components interconnecting the inner shaft 18 and the frame 28.The ball bearing 62 and the roller bearing 52 may be described as beingsupported by the frame 28 “in series” relative to radial loads. That is,the roller bearing 52 is mounted to the frame 28 through the firstbearing cage 32, and the ball bearing 62 is mounted to the secondbearing cage 44 which is mounted to the first bearing cage 32.

Because the second bearing cage 44 is a spring element, the radialcompressive load applied to the ball bearing 62 has a known relationshipto the relative radial deflection of the forward and aft rings 48 and 46of the second bearing cage 44 (i.e., the radial load on the ball bearing62 is proportional to the bending deflection of the spring arms 50).Considering a deflection of the inner shaft 18 towards the frame 28, theball bearing 62 will experience radial load only if there is relativedeflection between the ball bearing 62 and the roller bearing 52.

Given that the inner shaft 18 is a substantially rigid element, anyradial deflection of the inner shaft 18 will generally result in equaldeflection of the ball and roller bearings 62 and 52 with little to nobending of the spring arms 50.

Unequal deflection of the ball and roller bearings 62 and 52, orindependent radial deflection of the ball bearing 62, may occur based onthe presence of different radial clearances or preloads, or differentradial baseline positions between the roller bearing 52 and the ballbearing 62.

However, once the limit of any independent radial ball bearing movementis reached, further radial deflection of the ball and roller bearings 62and 52 will be in unison, with bending of the spring arms 38 of thefirst bearing cage 32. The radial deflection of the second bearing cage44, and thus the radial load on the ball bearing 62 cannot increasefurther. In essence, the ball bearing 62 is isolated from all but asmall, selected magnitude of the total radial deflection and radialloads that the bearing pair experiences in operation. In order toachieve maximum isolation of radial loads from the ball bearing 62, itis preferred that the spring arms 50 should have the lowest radialstiffness possible while still reliably transmitting axial loads. Thetotal radial load on the ball bearing 62 may be set at an arbitrarilyselected limit by manipulating the radial stiffness of the spring arms50, and relative eccentricities of the rings 48 and 46. As an example,the total radial load may readily be limited to well under 0.4 kN (100lbs.).

As described above, the ball bearing 62 and the roller bearing 52 arecoaxial with each other in a static condition, or in other words theaxis of rotation of each bearing coincides with axis of rotation of theother bearing. However, in some applications it may be desirable toprovide a radial preload to the roller bearing 52 in order to preventroller skidding. Therefore the ball bearing 62 can be laterally offsetfrom the roller bearing 52. This may be accomplished, for example, bymaking the bore of the forward ring 48 of the second bearing cage 44slightly eccentric from the bore of the aft ring 46 of the secondbearing cage 44, or by offsetting the spring arms 50 so they do notextend exactly parallel to the longitudinal axis 11. Under the influenceof thrust loads during engine operation, the ball bearing 62 willattempt to center the inner shaft 18 on its own rotational axis whichwill exert a limited radial force on the ball and roller bearings 62 and52.

This technique is generally referred to as “offsetting” and is aneffective means of guaranteeing that a minimum roller bearing load ispresent to prevent skidding. The present invention permits the bearings62 and 52 to be offset from each other while maintaining the rollerbearing 52 coaxial with the longitudinal axis of the engine 11. This isnot possible with a prior art “parallel” bearing support configurationdesign because it would cause the roller bearing 52 to become uncenteredwithin the oil film damper apparatus, decreasing the dampingeffectiveness. When offsetting is implemented the radial load on theball bearing 62 will be greater than it would be without offsetting,however the radial load on the ball bearing may still be limited to avalue well within the limit needed for adequate service reliability, forexample the total radial load may be about 2 kN (500 lbs.) or less.

FIG. 3 illustrates an alternative bearing support apparatus for mountingbearings between a frame 28 having a flange 30 and an inner shaft 18. Abearing cage 132 is mounted to the flange 30. The bearing cage 132 canbe considered a spring support or resilient support. The bearing cage132 comprises an annular forward ring 134, an annular middle ring 136,and an annular aft ring 138. The forward ring 134 and the middle ring136 are interconnected by an annular array of forward spring arms 140.The middle ring 136 and the aft ring 138 are interconnected by anannular array of aft spring arms 142.

The forward spring arms 140 are configured to be essentially rigidagainst axial deflection of the middle ring 136 while permittingcontrolled radial deflection of the middle ring 136 relative to theforward ring 134. The number, shape, spacing, and dimensions of theforward spring arms 140 may be modified to suit a particularapplication, in particular to achieve a desired radial stiffness.

The aft spring arms 142 are configured to be essentially rigid againstaxial deflection of the aft ring 138 while permitting controlled radialdeflection of the aft ring 138 relative to the middle ring 136. Thenumber, shape, spacing and dimensions of the aft spring arms 142 may bemodified to suit a particular application, in particular to achieve adesired radial stiffness. In some applications the radial stiffness ofthe aft spring arms 142 may be selected to be significantly less thanthe radial stiffness of the forward spring arms 140.

The middle ring 136 and the aft ring 138 are each generally cylindricaland extend axially in-line with the spring arms 140 and 142. The forwardring 134 extends in the radial direction so as to define a mountingflange. The forward ring 134 also includes a plurality of mounting holes144 which receive fasteners 146 that secure the forward ring 134 to theflange 30 of the frame 28.

A roller bearing 152 is disposed between the frame 28 and the innershaft 18 and includes an annular inner race 154, a row of generallycylindrical rollers 156, an annular outer race 158, and a retainer 160.The inner race 154 is secured to the inner shaft 18 so that it will notrotate relative thereto, and the outer race 158 is secured to the middlering 136 so that it will not rotate relative thereto.

A ball bearing 162 is disposed between the frame 28 and the inner shaft18, axially aft of the roller bearing 152. The ball bearing 162 includesan annular inner race 164, a row of spherical balls 166, an annularouter race 168, and a retainer 170. The inner race 164 is secured to theinner shaft 18 so that it will not rotate relative thereto, and theouter race 168 is secured to the aft ring 138 so that it will not rotaterelative thereto. The ball bearing 162 is configured as described abovesuch that axial loads (also referred to as thrust loads) can betransmitted from the inner shaft 18 through the ball bearing 162 and thebearing cage 132 to the frame 28.

The bearing support apparatus includes an optional annular damper arm180. The damper arm 180 has a forward portion 182 which is secured tothe flange 30, for example using the illustrated bolts 184 or othersuitable fasteners. The damper arm 180 extends radially inward andaxially aft from the flange 30. An aft portion 185 of the damper arm 180is generally cylindrical and is mounted in axial alignment and closeradial proximity to the middle ring 136 of the first bearing cage 132.

Together the outer surface of the middle ring 136 and the inner surfaceof the aft portion 185 of the damper arm 180 define a thin annularsqueeze film space 186 therebetween. A damper fluid such as an oil underpressure, is introduced into squeeze film space 186 through anappropriate conduit or inlet (not shown). In accordance with knownprinciples, any rotor imbalance may cause the inner shaft 18 and theroller bearing 152 to undergo radial motion and subject oil in thesqueeze film space 186 to very high pressure, thereby forcing viscousflow of the oil and a damping action on the roller bearing 152 and innershaft 18. In order to seal damper oil in squeeze film space 186, themiddle ring 36 includes a pair of spaced apart grooves 188 in its outersurface. A pair of sealing rings (not shown) may be fitted in thegrooves 188 and adapted to engage the surface of the damper arm 180,thereby sealing the forward and aft boundaries of the squeeze film space186.

The operation and performance of the bearing apparatus shown in FIG. 3is substantially the same as described above, with the ball bearing 162being isolated from all but minor radial deflections and loads by way ofthe series mounting arrangement provided by the bearing cage 132. Theaft ring 138 may be offset relative to the middle ring 136 as describedabove to provide a desired radial load on the roller bearing 152.

FIG. 4 illustrates another alternative bearing support apparatus. Theapparatus is similar in operational principle to those described above,but differs in that it is configured for mounting bearings between tworotating shafts, for example the inner shaft 18 described above and theouter shaft 26. A bearing cage 232 is mounted to the inner shaft 18 sothat it rotates with the inner shaft 18, for example by an interferencefit. The bearing cage 232 can be considered a spring support orresilient support. The bearing cage 232 comprises an annular forwardring 234, an annular middle ring 236, and an annular aft ring 238. Theforward ring 234 and the middle ring 236 are interconnected by anannular array of forward spring arms 240. The middle ring 236 and theaft ring 238 are interconnected by an annular array of aft spring arms242.

The forward spring arms 240 are configured to be essentially rigidagainst axial deflection of the middle ring 236 while permittingcontrolled radial deflection of the middle ring 236 relative to theforward ring 234. The number, shape, spacing, and dimensions of theforward spring arms 240 may be modified to suit a particularapplication, in particular to achieve a desired radial stiffness.

The aft spring arms 242 are configured to be essentially rigid againstaxial deflection of the aft ring 238 while permitting controlled radialdeflection of the aft ring 238 relative to the middle ring 236. Thenumber, shape, spacing and dimensions of the aft spring arms 242 may bemodified to suit a particular application, in particular to achieve adesired radial stiffness. In some applications the radial stiffness ofthe aft spring arms 242 may be selected to be significantly less thanthe radial stiffness of the forward spring arms 240.

The middle ring 236 and the aft ring 238 are each generally cylindricaland extend axially in-line with the spring arms 240 and 242. The forwardring 234 extends radially inward to engage the inner shaft 18.

A roller bearing 252 is disposed between the inner shaft 18 and theouter shaft 26 and includes an annular inner race 254, a row ofgenerally cylindrical rollers 256, an annular outer race 258, and aretainer 260. The inner race 254 is integral with (or secured to) themiddle ring 236 so that it will not rotate relative thereto, and theouter race 258 is secured to the outer shaft 26 so that it will notrotate relative thereto.

A ball bearing 262 is disposed between the inner shaft 18 and the outershaft 26, axially aft of the roller bearing 252. The ball bearing 262includes an annular inner race 264, a row of spherical balls 266, anannular outer race 268, and a retainer 270. The inner race 264 issecured to the aft ring 238 so that it will not rotate relative thereto,and the outer race 268 is secured to the outer shaft 26 so that it willnot rotate relative thereto. The ball bearing 262 is configured asdescribed above such that axial loads (also referred to as thrust loads)can be transmitted from the inner shaft 18 through the bearing cage 232and the ball bearing 262 to the outer shaft 26.

The bearing support apparatus includes an optional annular damper ring280. The damper ring 280 is generally cylindrical and is mounted to theinner shaft 18 in axial alignment and close radial proximity to themiddle ring 236 of the bearing cage 232.

Together the inner surface of the middle ring 236 and the outer surfaceof the damper ring 280 define a thin annular squeeze film space 286therebetween. A damper fluid such as an oil under pressure, isintroduced into squeeze film space 286 through an appropriate conduit orinlet 287. In order to seal damper oil in squeeze film space 286, themiddle ring 236 includes a pair of spaced apart grooves 288 in its outersurface. A pair of sealing rings 289 may be fitted in the grooves 288and adapted to engage the surface of the damper ring 280.

The operation and performance of the bearing apparatus shown in FIG. 4is substantially the same as described above, with the ball bearing 262being isolated from all but minor radial deflections and loads by way ofthe series mounting arrangement provided by the bearing cage 232. Theaft ring 238 may be offset relative to the middle ring 236 as describedabove to provide a desired radial load on the roller bearing 252.

FIG. 5 illustrates another alternative bearing support apparatus that itis configured for mounting bearings between two rotating shafts, forexample the inner shaft 18 described above and the outer shaft 26. Theapparatus is similar in operational principle to the apparatus shown inFIG. 4, but differs in the arrangement in of the spring elements. Abearing cage 332 is mounted to the inner shaft 18 so that it rotateswith the inner shaft 18, for example by an interference fit, andincludes an annular forward ring 334, an annular middle ring 336, and anannular aft ring 338. The middle ring 336 and the aft ring 338 areinterconnected by an annular array of inner spring arms 340. The forwardring 334 and the aft ring 338 are interconnected by an annular array ofouter spring arms 342.

The inner spring arms 340 are configured to be essentially rigid againstaxial deflection of the aft ring 338 while permitting controlled radialdeflection of the aft ring 338 relative to the middle ring 336. Thenumber, shape, spacing, and dimensions of the inner spring arms 340 maybe modified to suit a particular application, in particular to achieve adesired radial stiffness.

The outer spring arms 342 are configured to be essentially rigid againstaxial deflection of the forward ring 334 while permitting controlledradial deflection of the forward ring 334 relative to the aft ring 338.The number, shape, spacing and dimensions of the outer spring arms 342may be modified to suit a particular application, in particular toachieve a desired radial stiffness. In some applications the radialstiffness of the outer spring arms 342 may be selected to besignificantly less than the radial stiffness of the inner spring arms340.

The forward ring 334 and the aft ring 338 are each generally cylindricaland extend axially parallel with the spring arms 340 and 342. The middlering 336 extends radially inward to engage the inner shaft 18.

A roller bearing 352 is disposed between the inner shaft 18 and theouter shaft 26 and includes an annular inner race 354, a row ofgenerally cylindrical rollers 356, an annular outer race 358, and aretainer 360. The inner race 354 is integral with (or secured to) theaft ring 338 so that it will not rotate relative thereto, and the outerrace 358 is secured to the outer shaft 26 so that it will not rotaterelative thereto.

A ball bearing 362 is disposed between the inner shaft 18 and the outershaft 26, axially forward of the roller bearing 352. The ball bearing362 includes an annular inner race 364, a row of spherical balls 366, anannular outer race 368, and a retainer 370. The inner race 364 issecured to the forward ring 334 so that it will not rotate relativethereto, and the outer race 368 is secured to the outer shaft 26 so thatit will not rotate relative thereto. The ball bearing 362 is configuredas described above such that axial loads (also referred to as thrustloads) can be transmitted from the inner shaft 18 through the bearingcage 332 and the ball bearing 362 to the outer shaft 26.

The bearing support apparatus includes an optional annular damper ring380. The damper ring 380 is generally cylindrical and is mounted to theinner shaft 18 in axial alignment and close radial proximity to the aftring 338 of the bearing cage 332.

Together the inner surface of the aft ring 338 and the outer surface ofthe damper ring 380 define a thin annular squeeze film space 386therebetween. A damper fluid such as an oil under pressure, isintroduced into squeeze film space 386 through an appropriate conduit orinlet 387. In order to seal damper oil in squeeze film space 386, theaft ring 338 includes a pair of spaced apart grooves 388 in its innersurface. A pair of sealing rings 389 may be fitted in the grooves 388and adapted to engage the surface of the damper ring 380.

The operation and performance of the bearing apparatus shown in FIG. 5is substantially the same as described above, with the ball bearing 362being isolated from all but minor radial deflections and loads by way ofthe series mounting arrangement provided by the bearing cage 332. Theforward ring 334 may be offset relative to the aft ring 338 as describedabove to provide a desired radial load on the roller bearing 352.

It is noted that generally any of the bearing races described herein maybe made integral to one of the rings of the bearing cage supporting thatbearing.

In comparison to prior art bearing supports, the apparatus describedherein has the advantage of isolating a ball bearing from radial loads.This simplifies its design, allowing internal geometry to be optimizedfor pure thrust loading, and will result in longer life and improvedreliability. At the same time, increasing the radial load on theadjacent roller bearing increases cage speed, and thereby decreases therisk of roller skidding damage. The present invention is also simplerand cheaper than prior art designs used to isolate ball bearings fromradial loads.

The foregoing has described a bearing support apparatus for a gasturbine engine. While specific embodiments of the present invention havebeen described, it will be apparent to those skilled in the art thatvarious modifications thereto can be made without departing from thespirit and scope of the invention. Accordingly, the foregoingdescription of the preferred embodiment of the invention and the bestmode for practicing the invention are provided for the purpose ofillustration only and not for the purpose of limitation, the inventionbeing defined by the claims.

What is claimed is:
 1. A bearing support apparatus for a gas turbineengine, comprising: a first annular member; a second annular memberdisposed within the first annular member; first and second axiallyspaced-apart rolling-element bearings disposed between the first andsecond annular members so as to enable relative rotation of the firstand second members; a first spring element mounting the first bearing tothe first member; and a second spring element mounting the secondbearing to the first member through the first spring element, such thatthe first and second bearings are physically supported in seriesrelative to the first member.
 2. The bearing support apparatus of claim1 wherein the first bearing is a roller bearing.
 3. The bearing supportapparatus of claim 1 wherein the second bearing is a ball bearing. 4.The bearing support apparatus of claim 1 wherein the first bearing islaterally offset relative to the second bearing.
 5. The bearing supportapparatus of claim 1 wherein the second bearing is configured totransmit axial loads between the first and second members.
 6. Thebearing support apparatus of claim 1 wherein the first spring element isa first bearing cage mounted to first member and carrying the firstbearing.
 7. The bearing support apparatus of claim 6 wherein the firstbearing cage comprises annular forward and aft rings interconnected byannular array of axially-extending spring fingers.
 8. The bearingsupport apparatus of claim 1 wherein the second spring element is a cagemounted to the first spring element and carrying the second bearing. 9.The bearing support apparatus of claim 8 wherein the second bearing cagecomprises annular forward and aft rings interconnected by annular arrayof axially-extending spring fingers.
 10. The bearing support apparatusof claim 1 wherein the first and second spring elements are arranged inan axially overlapping relationship.
 11. The bearing support apparatusof claim 1 wherein the first and second spring elements are integralportions of a single member having annular forward, middle, and aftrings, the forward and middle rings being interconnected by annulararray of axially-extending forward spring fingers and the middle and aftrings being interconnected by annular array of axially-extending aftspring fingers.
 12. The bearing support apparatus of claim 1 furtherincluding a squeeze film damper disposed between the first member andone of the spring elements.
 13. The bearing support apparatus of claim 1wherein the first spring element has a first radial stiffness, and thesecond spring element has a second radial stiffness which issubstantially less than the first radial stiffness.
 14. The bearingsupport apparatus of claim 1 wherein the first and second springelements are separate members which are mechanically coupled together.15. The bearing support apparatus of claim 1 wherein the first member isa stationary frame.
 16. The bearing support apparatus of claim 14wherein the second member is a shaft mounted for rotation in the frame.17. A bearing apparatus for a gas turbine engine, comprising: astationary, annular frame; a first bearing cage comprising annularforward and aft rings interconnected by an annular array ofaxially-extending first spring fingers, wherein the forward ring ismounted to the frame; a second bearing cage comprising annular forwardand aft rings interconnected by an annular array of axially-extendingsecond spring fingers, the aft ring mounted to the aft ring of the firstbearing cage; a roller bearing carried by the aft ring of the secondbearing cage; a ball bearing carried by the forward ring of the secondbearing cage; and a shaft mounted in the roller and ball bearings. 18.The bearing apparatus of claim 17 wherein the ball bearing is configuredto transmit axial loads between the frame and the shaft.
 19. The bearingapparatus of claim 17 wherein the first and second bearing cages arearranged in an axially overlapping relationship.
 20. The bearing supportapparatus of claim 17 wherein the ball bearing is laterally offsetrelative to the roller bearing.
 21. The bearing apparatus of claim 17further including a squeeze film damper disposed between the frame andthe first spring cage.
 22. The bearing apparatus of claim 17 wherein thefirst bearing cage has a first radial stiffness, and the second bearingcage has a second radial stiffness which is substantially less than thefirst radial stiffness.